1. Field of Invention
This invention relates to hydrodynamic rotary seals that are used to retain lubricant and exclude the environment in diverse applications. More specifically, this invention relates to features that improve seal cross-sectional stability and exclusion edge circularity in zero and low differential pressure conditions, and conditions of low level reversing differential pressure, and provide improved contact pressure control at the dynamic sealing interface for improved abrasive exclusion, and improved consistency of hydrodynamic lubrication and flushing action.
2. Description of Related Prior Art
The following commonly assigned patent documents are related to the invention, and are incorporated herein by reference for all purposes:
United States Patents:
    1. U.S. Pat. No. 7,052,020 Hydrodynamic Rotary Seal    2. U.S. Pat. No. 6,767,016 Hydrodynamic Rotary Seal With Opposed Tapering Seal Lips    3. U.S. Pat. No. 6,685,194 Hydrodynamic Rotary Seal With Varying Slope    4. U.S. Pat. No. 6,561,520 Hydrodynamic Rotary Coupling Seal    5. U.S. Pat. No. 6,494,462 Rotary Seal With Improved Dynamic Interface    6. U.S. Pat. No. 6,382,634 Hydrodynamic Seal With Improved Extrusion Abrasion and Twist Resistance    7. U.S. Pat. No. 6,334,619 Hydrodynamic Packing Assembly    8. U.S. Pat. No. 6,315,302 Skew Resisting Hydrodynamic Seal    9. U.S. Pat. No. 6,227,547 High Pressure Rotary Shaft Sealing Mechanism    10. U.S. Pat. No. 6,120,036 Extrusion Resistant Hydrodynamically Lubricated Rotary Shaft Seal    11. U.S. Pat. No. 6,109,618 Rotary Seal With Enhanced Lubrication and Contaminant Flushing    12. U.S. Pat. No. 6,036,192 Skew and Twist Resistant Hydrodynamic Rotary Shaft Seal    13. U.S. Pat. No. 6,007,105 Swivel Seal Assembly    14. U.S. Pat. No. 5,873,576 Skew and Twist Resistant Hydrodynamic Rotary Shaft Seal    15. U.S. Pat. No. 5,823,541 Rod Seal Cartridge for Progressing Cavity Artificial Lift Pumps    16. U.S. Pat. No. 5,738,358 Extrusion Resistant Hydrodynamically Lubricated Multiple Modulus Rotary Shaft Seal    17. U.S. Pat. No. 5,678,829 Hydrodynamically Lubricated Rotary Shaft Seal With Environmental Side Groove    18. U.S. Pat. No. 5,230,520 Hydrodynamically Lubricated Rotary Shaft Seal Having Twist Resistant Geometry    19. U.S. Pat. No. 5,195,754 Laterally Translating Seal Carrier For a Drilling Mud Motor Sealed Bearing Assembly    20. U.S. Pat. No. 4,610,319 Hydrodynamic Lubricant Seal For Drill BitsUnited States Patent Applications:    1. Pub. No. 2005/0093246 Rotary Shaft Sealing Assembly    2. Pub. No. 2006/0214379 Composite, High Temperature, Dynamic Seal and Method of Making Same    3. Pub. No. 2006/0214380 Low Torque Hydrodynamic Lip Geometry for Bi-Directional Rotation Seals    4. Ser. No. 11/488,746 Filled Hydrodynamic Seal With Contact Pressure Control, Anti-Rotation Means and Filler Retention Means
Kalsi Engineering manufactures various configurations of hydrodynamic seals, based on the above-referenced patents and patent applications, and sells them under the registered trademark “KALSI SEALS.” The rotary seals that are marketed by Kalsi Engineering are installed with radial interference (i.e., compression), and seal by blocking the leak path. These well-known seals employ various dynamic lip geometries that cause a lubricant-side edge of a dynamic sealing interfacial contact footprint to be wavy. For example, see FIG. 13 of U.S. Pat. No. 5,230,520, FIG. 2F of U.S. Pat. No. 6,109,618, and FIGS. 2, 2A and 2B of U.S. Patent Application Publication No. 2006/0214380. As a consequence of the wavy lubricant-side footprint edge, the rotary motion of the lubricant-wetted shaft drags lubricant into the dynamic sealing interface, and causes the seal to hydroplane on a film of lubricant that separates the seal from the shaft. This hydrodynamic operating regime allows the seal to operate cooler and with less wear, even under conditions of high differential pressure acting from the lubricant side of the seal. The environment side of the interfacial contact footprint is intended to be circular rather than wavy, to avoid hydrodynamic activity with the environment, and thereby exclude the environment-but in fact the environment side of the footprint is typically quite wavy in prior art non-axially constrained seals that are used in conditions of low or no differential pressure, as described below. In the preceding sentence, the word “circular” is meant to imply that the environment side of the interfacial contact is in theory intended to form a circumference as it is described and understood in plane geometry; i.e. “a closed plane curved line with all of its points equidistant from an interior and coplanar point which is called the center” (Lacret Plane Geometry, March 1982). Although this environment-side circularity is desirable in theory, true perfect theoretical circularity is seldom if ever obtainable in any feature of any manufactured product in actual practice.
Virtually all of the above-noted commercial seals employ the projecting fixed-width annular static sealing lip as shown and described in FIG. 3 of U.S. Pat. No. 5,230,520. This fixed-width static lip provides an approximation of compressive symmetry, and provides a dramatic increase in abrasive exclusion in low differential pressure conditions, compared to the original first-generation seals that were based on U.S. Pat. No. 4,610,319 and did not employ a projecting static lip.
Head-to-head testing of seals with and without such fixed-width projecting static lips was performed, and the abrasive exclusion performance of the seals with the fixed-width static sealing lip were very superior to the first generation seals in low differential pressure conditions. While providing a clear and consistent benefit in low differential pressure conditions, the abrasive exclusion performance of the fixed-width static lip was not consistent in zero differential pressure conditions.
In zero differential pressure conditions in the presence of an abrasive environment, some seals with the fixed width static lip performed well in the laboratory for hundreds of hours, while others lasted only a few hours before third-body abrasion took its toll. This zero differential pressure problem was initially addressed by axially spring loading the seals from the lubricant side to simulate low differential pressure, as shown in FIGS. 3-8 of the Kalsi Seals Handbook, Rev. 1. Later, special purpose seals were developed that employed the seal body as an axial spring, as disclosed in U.S. Pat. No. 6,315,302. These “axially constrained” seals have been very successful in applications where a lubricant must be partitioned from an abrasive environment in conditions of zero differential pressure, or low levels of reversing differential pressure. Such axially constrained seals presently have limitations in terms of their ability to handle high differential pressure acting from the lubricant side. Also, it has been noted that the exclusion edge chamfer of such axially constrained seal is subject to flattening in high differential pressure conditions.
A seal is needed that is capable of handling high differential pressure, while offering good abrasion resistance in conditions of periodic zero differential pressure conditions.
Rotary seals with smaller radial cross-sections, to fit smaller radial groove depths, are desirable in oilfield mud motor sealed bearing assemblies. The smaller radial groove depth means that for any given size of motor, the wall thickness of the shaft and housing can be increased, for improved strength. This is particularly important in miniature mud motor sealed bearing assemblies.
For the same amount of radial dimensional compression, the percentage of compression of a small seal cross-section is greater than that of a larger seal cross-section, which means that interfacial contact pressure increases as the seal cross-sectional size decreases. For example, a nominal radial dimensional compression of 0.030″ results in 10% nominal compression with a 0.30″ deep seal cross-section, and results in 20% nominal compression with a 0.15″ deep radial seal cross-section. Since in general terms, the interfacial contact pressure is related to the percentage of compression times the modulus of elasticity of the seal material, the smaller cross-section seal has significantly higher interfacial contact pressure than the larger cross-section seal. This increased interfacial contact pressure can make the small cross-section seal more difficult to lubricate.
For the same amount of radial dimensional compression, smaller diameter seals experience a higher percentage of circumferential compression than larger diameter seals, and this effect also increases the interfacial contact pressure of smaller diameter seals.
In summary, interfacial contact pressure increases as a function of decreasing the seal cross-section and/or decreasing the seal diameter. This increasing contact pressure effect associated with small seal cross-sections and diameters is magnified by the exclusion edge chamfer used on existing commercial axially constrained seals. The use of an exclusion edge chamfer to manage interfacial contact pressure near the exclusion edge of the seal therefore becomes less practical as the cross-sectional size and/or diameter of an axially constrained seal is miniaturized. An alternate method of controlling interfacial contact pressure near the exclusion edge is desirable for axially constrained seals having a small cross-section and/or diameter.
The fixed-width projecting static lip was originally designed using finite element analysis, at a time when the available analytical tools (e.g. software and computers) were relatively primitive. Element choice was limited, and less than ideal. Mesh size was limited by computer processing power, and run-times were extremely long, even with two dimensional axi-symmetric models. The insights gained were sufficient to realize improved seal performance through the use of the fixed-width projecting static lip, but subtleties were masked by the limited analysis technology of the period.
Dramatic computer and software advances now permit three dimensional finite element analysis with advanced element types using highly refined meshes that provide a more detailed understanding of seal performance. This analysis reveals that in non-axially constrained seals exposed to zero differential pressure conditions, the fixed-width static lip only provides optimum cross-sectional stability near the average width of the dynamic lip. At some circumferential locations, the static lip twists away from the environment and causes the exclusion edge to lift away from the shaft in a wavy pattern at a low angle of convergence that is responsible for wedging the abrasive environment into the dynamic sealing interface. This undesirable seal attitude in zero-differential pressure conditions has been noted both with single modulus seals constructed in accordance with U.S. Pat. No. 5,230,520, and with dual modulus seals constructed in accordance with U.S. Pat. No. 5,738,358.
The changing cross-sectional twist between the narrowest and widest portions of the dynamic lip cause the seal to be significantly non-circular in zero differential pressure conditions, pre-disposing it to significant skewing as a result of circumferential compression and circumferential differential thermal expansion between the seal and the seal gland, resulting in skew-induced abrasive ingestion.
Finite element analysis shows that these undesirable seal characteristics are rectified by axial spring loading from the lubricant side of the seal, or by employing low differential pressure from the lubricant side, thus helping to explain performance observed in the laboratory. Seal to gland friction is present in real world conditions. Sometimes the friction is sufficient to retain the seal in a suitable orientation for abrasive exclusion, and sometimes it is not, resulting in wide variations in abrasive exclusion that have been observed from test to test in zero differential pressure conditions.
The fixed-width static lip of Kalsi Engineering's “High Film” seal, disclosed in U.S. Pat. No. 6,109,618, was initially developed intuitively since, at the time it was developed, suitably efficient three dimensional modeling and analysis techniques were not available. The resulting seals had wide variations in hydrodynamic flushing action, even in conditions of differential pressure acting from the lubricant side. For example, if a seal was tested and then disassembled for observation, then reinstalled and retested, it might have dramatically different lubricant flushing in the first and second portions of the test. This problem was ultimately rectified by experimentally increasing the width of the static lip. This experimental work provided the new insight that the width of the static lip plays a definite role in lubricating and flushing efficiency and consistency by providing improved uniformity in contact pressure near the lubricant side of the interfacial contact footprint. Unfortunately, the static lip fixed-width that provides the best uniformity in trailing edge contact pressure from one assembly to another is different than the static lip width that provides the best condition of average compressive symmetry over the varying width of the dynamic sealing interface for achieving abrasive exclusion.
Seals have been proposed that have opposed hydrodynamic lips (see, for example, FIGS. 8 and 8A in U.S. Pat. No. 6,685,194), where the seal is allowed to slip rotationally with respect to both the shaft and the gland. In such seals, good compressive symmetry could be achieved at every circumferential location for improved exclusion edge contact pressure and circularity. Such seals are not appropriate for conditions of high differential pressure, because if a rotary seal is allowed to slip very much with respect to the extrusion gap, the sharp extrusion gap corner of the gland can cut the seal and cause it to fail prematurely, especially if any corner defects are present. While seals that have opposed hydrodynamic lips may be suitable for zero differential pressure or low levels of reversing differential pressure, they are not considered suitable for high differential pressure service.
Slippage of the seal within the housing gland can also occur with seals that have a fixed-width static lip. This occurs more often in low differential pressure service with large diameter seals because the moment arms between the static and dynamic interfaces are more nearly equal. Slippage of seals with fixed-width static lips has also been observed in high differential pressure service, when the seal is exposed to reversing pressure differential. Rotational slippage is particularly undesirable in large diameter seals. The slippage can vary around the circumference of a large seal, causing undesirable localized circumferential stretching. Slippage is exacerbated by wear of the dynamic lip and/or the mating shaft surface, because such wear causes an increase in seal running torque.
Seals having anti-rotation projections molded into the lubricant end of the seal to engage recesses in the lubricant-side gland wall are disclosed in U.S. Application Pub. No. 2005/0093246 A1. Such seals are suitable for conditions of constant differential pressure from the lubricant side, or zero differential pressure. Such seals are not suitable for differential pressure acting from the environment side, because even low differential pressure from the environment side causes the seal to bow into the mating recesses in the lubricant-side gland wall, resulting in skewing of the exclusion edge of the seal, which promotes skew-induced environmental abrasion of the seal.
It has recently been observed by the inventors that with seals having a variable dimension dynamic lip, the length of the compressed seal varies as the width of the dynamic lip varies around the circumference of the seal. The compressed seal body is longest near the widest part of the dynamic lip, and is shortest near the narrowest part of the dynamic lip. This means that if the seal is forced substantially flat against the lubricant side gland wall by differential pressure acting from the environment side, the environment end of the seal will become wavy owing to the compressed length variations of the seal body. This environment end waviness has negative implications in terms of environmental exclusion, because it promotes environmental abrasion of the seal.
It is desirable to be able to overcome the shortcomings described above. A seal is needed that is suitable for high differential pressure acting from the lubricant side, while offering improved abrasion resistance performance in the periodic conditions of zero differential pressure or low levels of reversing pressure that are common to many abrasive environment applications.